Heat exchanger having winding channels

ABSTRACT

A winding channel heat exchanger includes a heat transfer member having winding channels, a manifold, and a cover plate. The channels&#39; winding design is defined by a non-linear flow axis that may include a plurality of short pitch and small amplitude undulations, which cause the flow to change directions, and may also or alternatively include two or more large amplitude bends that cause the flow to reverse direction. In one embodiment, the undulations have varying amplitudes to increase the heat transfer coefficient along the length of the channel. The winding channels allow a user to customize the pressure drop to promote good flow distribution, to achieve improved heat transfer uniformity, and to improve the heat transfer coefficient.

RELATED APPLICATIONS

The present application claims priority under 35 USC §119(e) to U.S. provisional application No. 61/347,949, filed on May 25, 2010, and is a continuation-in-part under 37 CFR §1.53(b) of U.S. application Ser. No. 12/188,859, filed Aug. 8, 2008, the entire contents of all the above applications are herein incorporated by reference in their entirety.

GOVERNMENT LICENSE RIGHT

The U.S. Government has a paid-up license in this invention and the right in limited circumstances to require the patent owner to license others on reasonable terms as provided for by the terms of Contract No. N65540-06-C-0015 awarded by the U.S. Navy.

TECHNICAL FIELD

This invention relates generally to an apparatus for cooling a heat-producing device and, more specifically, to a liquid cooled heat exchanger having winding, non-linear channels.

BACKGROUND

The use of heat exchangers for cooling a range of heat producing devices, for example, electronic devices is known in the art. Liquid cooled heat exchangers are generally characterized as having macro-channels, mini-channels, or micro-channels, depending on the size of the channels. The term ‘micro’ is applied to devices having the smallest hydraulic diameters, generally between ten to several hundred micrometers, while ‘mini’ refers to diameters on the order of about 0.5 mm to about 2 millimeters, and ‘macro’ channels are the largest in size, generally greater than about 2 millimeters. An example of a typical macro channel design is the conventional swaged-tube cold plate illustrated in FIG. 1 a.

As shown in FIG. 1 a, the prior art swaged-tube cold plate includes a copper tube 11 swaged into grooves machined in an aluminum plate 13. Swaged tubes are generally suitable for cooling large-area devices, particularly when cost is a factor, and/or when the cooling requirements do not require a very low thermal resistance. The lowest thermal resistance that can generally be achieved with a conventional swaged-tube cold plate is approximately 2° C./(W/cm²). Because of these limitations, applications requiring lower thermal resistances often use finned cold plates, such as the prior art finned cold plate shown in FIG. 1 b.

Conventional finned cold plates have a number of closely spaced fins 21 attached to the heat transfer surface 23. The fluid flows through the channels 25 formed by the spaces between the fins. The channels typically have a width between about 1 to 5 mm. Conventional finned cold plates can achieve thermal resistances as low as approximately 1° C./(W/cm²).

The thermal resistance of macro channel cold plates decreases as the flow rate is increased and approaches asymptotically a minimum value at a flow of about 0.1 LPM/cm². Increasing the flow rate further has not been found to result in an additional reduction in the thermal resistance.

For cooling high heat flux devices, such as solid-state laser diodes, which dissipate heat at a rate of 500-1000 W/cm², cold plates with substantially lower thermal resistance than that of the swaged-tube cold plates or the machined fin cold plates are needed. In these applications, micro-channel cold plates are generally employed.

There are two primary types of prior art micro-channel cold plates: parallel flow and normal flow. As the name implies, parallel flow micro-channel cold plates have the liquid flowing through the heat transfer passages in a direction parallel to the surface being cooled. In contrast, normal flow micro-channel cold plates (NCP) have the liquid flowing through the heat transfer passages in direction normal to the surface being cooled. The parallel flow cold plates have geometries similar to that of the finned cold plate shown in FIG. 1 b, except that the dimensions are scaled down by an order of magnitude. For example, the channel width in a micro-channel cold plate is typically less than 500 microns. Because of the high-pressure drop in the micro-channels, the size of the parallel flow micro-channel cold plates is typically less than about 1 cm to 2 cm on a side. Even at these small sizes, the pressure drop can be too large for some applications. The pressure drop can be reduced by subdividing the micro-channel into several sections and providing alternating inlet and outlet manifolds along the length of the cold plate, for example as described in U.S. Pat. No. 6,986,382 to Upadhya.

One objective in the design of conventional micro-channel cold plates is to minimize the pressure drop consistent with achieving the target thermal performance. Minimizing the flow length and maximizing the flow area of the micro-channels is most often employed to achieve this objective. Conventionally, the flow length is minimized by making the flow axis straight, while making the micro-channel depth large compared to its width maximizes the flow area. As such, prior art parallel-flow micro-channels have a depth that is an order of magnitude larger than the width.

Normal flow cold plates invented by the present inventor, Javier Valenzuela and as described in U.S. Pat. Nos. 5,145,001 and 6,935,411 among other patents, demonstrate excellent heat transfer effectiveness, especially in high heat-flux applications. However, for some systems the highly effective cooling provided by the normal flow design is not required, and the cost of the heat exchanger may not be warranted. FIG. 2 shows one example of a cross-section of a prior-art normal flow micro-channel cold plate. Normal-flow micro-channel cold plates incorporate a low-pressure drop manifold structure 27 that distributes and collects the flow over the active area of the cold plate. The micro-channels 29 are embedded in a thin layer between the manifold structure and the active surface 31. The micro-channels direct the fluid in a direction substantially normal to the active surface: first from the manifold structure towards the active surface, and then from the active surface towards the manifold structure. The total length of the micro-channels is very short, about twice the thickness of the heat transfer layer and, therefore, the pressure drop in the normal flow micro-channel cold plates is small, even at the high flow rates per unit area required in these high heat flux applications.

In spite of the order of magnitude lower thermal resistances that can be obtained through the use of micro-channels, they are seldom used in large-area cold plates. The principal objections to the use of micro-channels in large area cold plates are: (1) the large pressure drop associated with the flow through long, small-hydraulic-diameter passages, and (2) the relatively high cost of fabricating passages with such small dimensions. The cost of fabrication and large pressure drops can also prevent micro-channels from being used in other applications as well. When micro-channels are not utilized, macro or mini-channels may be utilized and performance is sacrificed for cost, ease of use or other requirements.

There are also other methods of cooling that utilize fluid flowing through channels in order to cool a device. For example, U.S. Pat. No. 6,213,194 discloses the use of a hybrid cooling system for an electronic module which includes refrigeration cooled cold plate and an auxiliary air cooled heat sink. The '194 patent also discloses the use two independent fluid passages embedded in the same cold plate to provide redundancy. A single serpentine passage, akin to that of a swaged tube cold plate, or multiple straight passages feed by headers, akin to a finned cold plate, is used for each one of the redundant systems.

SUMMARY

In accordance with the present disclosure, there is provided a winding channel heat exchanger that includes a heat transfer member having winding channels for cooling a heat-producing device.

The channels' winding design is defined by a non-linear flow axis that, in one embodiment, has a plurality of short pitch and small amplitude undulations, which cause the flow of fluid in the channels to change directions. The winding channels may also include two or more large amplitude bends that cause the flow to reverse direction. The winding channels advantageously allow a user to customize the pressure drop to promote good flow distribution, achieve improved heat transfer uniformity, and increase the heat transfer coefficient.

The heat transfer member includes one or more heat transfer layers, each layer having one or more inlet openings and corresponding outlet openings. Each of the winding channels is in fluid communication with at least one of the inlet openings and at least one of the corresponding outlet openings, such that the cooling fluid enters the inlet openings, flows along the channels, and exits via the outlet openings. In one embodiment, the openings are arranged in rows through each layer, each opening extending from the first surface through to the second surface of each heat transfer layer.

A manifold can supply fluid to each of the inlet openings of the heat transfer member and receives fluid from each of the outlet openings of the heat transfer member. The manifold may distribute and collect the fluid throughout the active heat transfer area in order to promote uniform heat transfer throughout the area.

In alternate embodiments disclosed herein the configuration of the winding channels is modified according to the particular application, but in all embodiments, the winding channel's axis remains non-linear along at least a portion of the length of the channel.

BRIEF DESCRIPTION OF THE DRAWINGS

The foregoing and other objects, features and advantages will be apparent from the following description of particular embodiments of the invention, as illustrated in the accompanying drawings in which like reference characters refer to the same parts throughout the different views. The drawings are not necessarily to scale, emphasis instead being placed upon illustrating the principles of various embodiments of the invention.

FIG. 1 a is a perspective view of a prior art swaged-tube cold plate;

FIG. 1 b is a perspective view of a prior art finned cold plate with its cover raised;

FIG. 2 is a cross-sectional view of a prior art normal flow heat exchanger;

FIG. 3 a is an exploded perspective view of a winding channel heat exchanger according to a first embodiment of the present invention;

FIG. 3 b is an enlarged view of a winding channel of FIG. 3 a;

FIG. 4 a is a top plan view of the heat exchanger of FIG. 3 a, with its cover removed;

FIG. 4 b is a cross sectional view taken along lines 4 b-4 b of FIG. 4 a;

FIG. 5 is a comparison of the winding channel design according to the present application with prior art linear designs where:

FIG. 5 a is a schematic view of the winding channel of FIG. 3 b with numerals 0-5 representing the average fluid temperature in the winding channel segment according to the present invention; and

FIG. 5 b is a schematic view of a prior art linear channels with numerals 0-5 representing the average fluid temperature in the channels segment;

FIG. 6 is a comparison of the winding channel design according to the present application with prior art linear designs where:

FIG. 6 a is a schematic view of the winding channel of FIG. 3 b according to the present invention;

FIG. 6 b is a schematic view a second prior art linear channel configuration in which the distance between the inlet and exit openings remains the same as in the winding channel configuration of FIG. 6 a;

FIG. 7 is a partial top plan view in cross-section of an increasing amplitude mini-channel heat transfer member according to one embodiment of the present invention;

FIG. 7 a is a graph showing the heat transfer coefficient enhancement and pressure drop of exemplary winding channel geometry according to the present invention;

FIG. 7 b is a graph showing test data comparing performance of a prior art swage tube and normal flow heat exchangers to two exemplary winding channel heat exchangers according to the present invention;

FIG. 7 c is a graph comparing the pressure drop of a straight channel to that of two winding mini-channels each having different amplitude undulations according to the present invention;

FIG. 7 d is a graph comparing the pressure drop of a straight channel to that of two winding mini-channels each having different wavelength undulations according to the present invention;

FIG. 7 e is a graph comparing the heat transfer coefficient of a straight channel to that of two winding mini-channels having different amplitude undulations according to the present invention;

FIG. 7 f is a graph comparing the heat transfer coefficient of a straight channel to that of two winding mini-channels having different wavelength undulations according to the present invention;

FIG. 8 is an exploded perspective view of the winding channel heat exchanger of FIG. 3 a including a flow restrictor plate.

FIGS. 9 a-9 e are schematic views of alternate winding channel designs according to the present invention; and

FIG. 10 is a top plan view of an alternate embodiment of a heat transfer member including winding channels according to the present invention.

DETAILED DESCRIPTION

The embodiments disclosed herein relate to a heat exchanger having winding channels. As used herein, the term “winding” is used to mean a twisting, serpentine, sinuous path, or the like, which may have a curvature or be angular, and which creates a non-linear path between an inlet and an outlet. As also used herein the term micro-channel is used in the conventional sense with respect to liquid-cooled heat exchanger technology and does not have specific dimensional constraints, but is generally understood to mean channels having the smallest hydraulic diameters, generally between ten to several hundred micrometers, although it is understood that as industry standards change, so too may the dimensions of micro-channels. Likewise, as also used herein the term mini-channel is used in the conventional sense with respect to liquid-cooled heat exchanger technology, and does not have specific dimensional constraints, but is generally understood to mean channels that have diameters on the order of about 0.5 mm to about 2 millimeters, although it is understood that as industry standards change, so too may the dimensions of mini-channels.

Referring initially to FIGS. 3 a-4 b, a first, exemplary winding channel heat exchanger 10 including a manifold 12, a heat transfer member 14 having winding channels 30, and a cover plate 15, is illustrated. In use, heat is transferred to or from the heat exchanger 10 over the portion that is enclosed by the dashed line 17, otherwise referred to as the “active area”, which corresponds to the portion of the heat exchanger 10 that includes winding channels 30, as described in more detail below. During use, the heat transfer member 14 is placed into thermal contact with a device to be cooled, as is known in the art, and fluid flows through the winding channels 30 in order to cool the device. In the heat exchanger 10 of the present embodiment, the winding channels are illustrated as winding micro-channels formed in one or more layers 28 of heat transfer member 14. Alternately, the channels may be larger, mini-channels and may be formed in a single layer or multiple layers. The description that follows, while referring to micro-channels is not limited to such channels, and applies to other size channels as well, in particular mini-channels. The plurality of channels may be identical to the exemplary winding channel, as in the present embodiment, or may be varied. For example, the plurality of channels formed in the one or more layers may be mirror images, may be symmetrical or non-symmetrical, or may have different geometries from each other, and may be varied from the exemplary winding channel described herein, as described in greater detail below.

The winding micro-channels 30 according to the present application each include a non-linear flow axis 36, as best shown in FIG. 3 b. The non-linear flow axis 36 may include one or more undulations 38 that cause the flow to change directions, as well as one or more pairs of bends 40 a, 40 b that cause the flow to reverse direction. A reference line “CL” bisects the length of the channel 30 through the approximate center between the inlet opening 32 and outlet opening 34. In the following description, the portion of the channel 30 disposed between the line “CL” and the inlet opening 32 is referred to as the inlet side (I_(S)), and the portion of the channel disposed between the line “CL” and the outlet opening 34 is referred to as the outlet side (O_(S)). It will be understood that the line “CL” is provided for reference purposes only and is not part of the actual heat exchanger. The fluid flow is reversed in that the first bend 40 a reverses the direction of the fluid flow from traveling from the inlet side (I_(S)) toward the outlet side as represented by arrow A, to a direction traveling from the outlet side (O_(S)) toward the inlet side of the channel as represented by arrow B. Likewise, the second bend 40 b reverses the direction of the fluid that is now flowing toward the inlet side (arrow B), and re-directs the fluid flow back toward the outlet side (arrow C) of the channel. In order that the flow of fluid ultimately reaches the outlet side and outlet opening 34, for each bend that changes the direction of flow toward the inlet side I_(S) of the channel there is preferably a corresponding bend that changes the flow back toward the outlet side (O_(S)) of the channel.

In addition to the one or more pair of reversing bends 40 a, 40 b, the winding micro-channel 30 may also include one or more undulations 38 that change the direction of the fluid flow, but which do not reverse the direction of the fluid flow. In the present embodiment, the undulations 38 have a smaller amplitude “a” than that of the bends 40 a, 40 b. These smaller amplitude undulations 38 change the local direction of the fluid flow without reversing the overall direction so that the flow continues in the same overall direction the fluid was traveling before reaching the undulation. It will be appreciated that the number and size of the bends and undulations can be varied depending upon the particular application, and the micro-channels may include both bends and undulations or include only bends or only undulations.

In the present embodiment, a bonded stack of three layers 28 is illustrated and may be varied according the needs of the particular application, as would be known to those of skill in the art. Each one or more layer 28 is generally planar and includes a first surface 10 a and a second surface 10 b, opposite the first surface. The micro-channels 30 may be formed in the second surface, for example by etching, such that the micro-channels have a depth that is less than the thickness “t” of their corresponding layer. The micro-channels may be formed by alternate methods as would be known to those of skill in the art, and may have a depth equal to the corresponding layer in some embodiments. Unlike the prior art, parallel flow micro-channels; the winding micro-channels 30 of the present invention are characterized by having an aspect ratio close to unity (i.e., the depth of the micro-channels is comparable to the width). The depth of the micro-channels may preferably be between about ½ the size of that of the width to about 1½ the size of that of the width for the depth and width to be comparable. For example, if the width of the channel is 2 mm, the depth may be from about 1 mm to about 3 mm.

As best shown in FIG. 4 a, the heat transfer layers 28 each have a plurality of inlet openings 32 and corresponding outlet openings 34 arranged in substantially parallel rows a, b, c, d, e, and f through each layer, each opening extending from the first surface 10 a through to the second surface 10 b. Each winding micro-channel 30 is in fluid communication with at least one of the inlet openings 32 and at least one of the corresponding outlet openings 34. Each of the inlet openings 32 and outlet openings 34, in turn, is in fluid communication with corresponding inlet channels 22 and outlet channels 24 of the manifold 12 (FIG. 4 b). Each winding micro-channel 30 may share their inlet openings and/or outlet openings, although the winding micro-channels may alternately have independent inlet openings and outlet openings. During use, fluid is provided from the manifold 12 and flows into the inlet openings 32 and into each of the micro-channels 30. The fluid then flows through the micro-channels 30 and out of each of the outlet openings 34 which return the fluid to the manifold 12.

In the present embodiment, the functions of distributing and collecting the fluid over the active heat transfer area 17, and transferring the heat between the fluid and the active heat transfer area 17 may be achieved by two separate components: the manifold 12 and the heat transfer member 14, respectively. This separation in functions allows the selection of the flow passage geometry in each component to the benefit of their respective functions.

The manifold 12 may distribute and collect the fluid over the entire heat transfer surface 17 in order to promote uniform heat transfer over the surface. In the present embodiment, the manifold has an interdigitated design as described below. However, alternate manifold designs may be utilized, such as traditional linear manifolds that are not interdigitated and which may include only a single channel. As best shown in FIGS. 3 a and 4 a, in the interdigitated manifold, the fluid enters the heat exchanger 10 through inlet port 16 of the manifold, which may be disposed near a first edge of the manifold 12. Inlet port 16 is fluidly connected to the inlet header 20, such that fluid is fed from the inlet port 16 to the header 20. The header 20 distributes the inlet fluid along the y-axis of the manifold and is fluidly connected to the inlet channels 22, such that the fluid is fed from the header 20 to the inlet channels 22 which distribute the fluid along the x-axis of the manifold. The fluid is then directed to the heat transfer member 14 where it enters inlet openings 32, is directed through fluidly connected winding micro-channels 30, and exits the micro-channels through corresponding outlet openings 34. The fluid then flows into a plurality of outlet channels 24 that are interdigitated, i.e. alternating, with the inlet channels 22, along the x-axis of the manifold 12. The outlet header 26, in turn, collects the fluid exiting the outlet channels 24 along the y-axis of the manifold. Fluid exits the heat exchanger 10 through outlet port 18 that is disposed near the opposite edge of the manifold in the present embodiment.

In the winding micro-channel heat exchanger 10, as the fluid is being distributed and collected it is desirable to minimize the pressure drop in the manifold to promote good flow distribution. As discussed in greater detail below, it is also desirable to keep a suitably small separation between the heat transfer member inlet openings 32 and the respective outlet openings 34 to promote uniform heat transfer over the heat transfer area 17. These requirements are conflicting since a small pressure drop would favor large dimensions for the manifold channels 22 and 24, whereas a small separation between the heat transfer member inlet and outlet openings would favor small dimensions for the channels.

The distance between the inlet 32 and outlet openings 34 of the heat transfer member 14 determines the minimum length, and thereby the minimum pressure drop, of the micro-channel passages; and the distance also determines the degree of temperature uniformity (or heat transfer uniformity) which can be achieved throughout the heat transfer member. To make best use of the flow heat transport capacity, and thereby minimize the flow and pressure drop requirements for a given application, it is desirable that the fluid exit temperature be close to the temperature of the surface of the heat transfer member (i.e. high heat exchanger effectiveness). The temperature difference between the fluid and the micro-channel walls is greater near the inlet openings 32 than the outlet openings 34, thereby providing greater heat transfer capability near the inlet openings than near the outlet openings. The variation in heat transfer capability is mitigated by heat conduction in the heat transfer member 14 along a plane parallel to the heat transfer surface. For high thermal conductivity materials, such as copper, this mitigation is most effective when the distance between the inlet and outlet openings is no more than a factor of 5 to 10 times larger than the thickness of the heat transfer member 14. Hence for a member 0.5 mm thick, the distance between inlet and outlet ports should be between about 2.5 to 5 mm.

In some applications it may be desirable to increase the pressure drop if it is unacceptably low. For example, in large area cold plates (larger than about 2×2 cm), it has been determined that at certain flow rates, air bubbles can block linear micro-channels. It has been determined that at the intended water flow rate, the pressure drop through the micro-channel heat transfer member was lower than the bubble point of the micro-channels, and hence insufficient to drive the bubbles out of the micro-channels, an unexpected result of the use of the linear micro-channels in a large-area application. As such, any gas present in the system could block areas of the heat exchanger, resulting in undesirable hot spots. Moreover, since good flow distribution requires that the pressure drop in the manifold be an order of magnitude smaller that in the heat transfer member, such a low pressure drop in the heat transfer member would place undue constraints on the manifold pressure drop, requiring the use of much larger manifold. Thus, contrary to expectations and to the common perception that micro-channels are not desirable because they have too large a pressure drop, the Applicant determined that particularly in large-area applications the opposite was actually true. In particular, that the desired inlet-to-outlet channel spacing, the pressure drop of conventional micro-channels is too low at the typical flow rates employed in large-area cold plates to achieve acceptable performance.

As discussed in greater detail below, the winding channel configuration disclosed herein provides a way to regulate the pressure drop in the heat transfer member to a desired value, while at the same time improving both the heat transfer capability and the heat transfer uniformity of the heat exchanger. At the flow rates per unit area typical of large area cold plates, the flow in the micro-channels is laminar (Reynolds number typically less than about 500) and the pressure drop in the micro-channels is proportional to the product of the velocity and the micro-channel length, and inversely proportional to the micro-channel hydraulic diameter. Therefore, increasing the length, increasing the velocity, or decreasing the diameter can all increase the pressure drop. The performance of micro-channel heat exchangers improves as the diameter of the micro-channels is decreased. Hence the diameter is often selected as the minimum diameter consistent with other considerations, such as ease of manufacture, or filtration level requirements. Therefore, for the purpose of comparing different micro-channel configurations, the diameter is not considered a design variable.

FIGS. 5 and 6 compare two channel configurations (which could be micro or mini) designed to have nominally the same pressure drop when operating at the same flow per unit area of the cold plate. FIGS. 5 a and 6 a illustrate schematic diagrams including winding channels according to the present invention. FIG. 5 b illustrates a first, conventional, prior-art linear channel configuration in which the desired pressure drop has been achieved by making the overall length of the linear channel the same as that of the winding channel. Equal length is achieved by increasing the distance between the inlet and the exit openings. FIG. 6 b illustrates a second conventional, linear channel configuration in which the distance between the inlet and exit openings remains the same as in the winding channel configuration, and desired pressure drop has been achieved by increasing the flow velocity in the linear channel. Increased velocity is achieved by decreasing the number of channels per unit area of the cold plate.

As discussed below, the winding channel configuration is advantageous relative to the straight (i.e. linear) channel configurations on two counts: (1) improved heat transfer uniformity; and (2), greater average heat transfer capability. To illustrate the two points, the total length of each channel depicted in FIGS. 5 and 6 has been divided into 6 segments of equal length. The numerals 0-5 are assigned to each segment to represent the average fluid temperature in that segment, the lowest number, “0”, near the inlet opening representing the coolest temperature, and the highest number, “5”, near the exit opening representing the highest temperature. The heat transfer surface in FIG. 5 is further divided by the dashed lines into six vertical columns, or strips, labeled A-F. Similarly, the heat transfer surface in FIGS. 6 a and 6 b is divided by the dashed lines into two vertical columns, or strips, labeled A-B and A′-B′. For each channel configuration in FIGS. 5 and 6, the average fluid temperature for each strip is then computed as numerical average of the fluid temperatures within that strip. The average strip temperatures are indicated with an underscore.

For the winding configurations 5 a and 6 a the average fluid temperatures alternate between values of 2.3 to 2.7 between the strips. For the linear channel configuration 5 b with increased distance between the inlet and exit ports, the average fluid temperatures range from 0 near the inlet port to 5 near the exit port. For linear channel configuration 6 b with increased velocity in the channels, the average fluid temperatures alternate between 1 and 4. It will be readily appreciated that the winding configuration disclosed in the present invention aids in providing greater uniformity in fluid temperature, and hence greater uniformity in heat transfer, over the heat transfer surface.

The winding channels of FIGS. 5 a and 6 a are also advantageous over the prior art linear channels because they provide a higher average heat transfer capability. The average heat transfer capability of the heat transfer member is a function of the product of the total channel wall area and the channel heat transfer coefficient: the higher this product, the higher the average heat transfer capability of the heat exchanger. For a conventional, linear micro-channel operating in the laminar flow regime, the heat transfer coefficient depends only on the micro-channel geometry and fluid properties, and is independent of the fluid velocity. For the winding channels, in contrast, the frequent changes in flow direction result in a heat transfer coefficient larger than that of a linear channel of equal cross-section, and the magnitude of the enhancement increases with increasing velocity. Hence, the average heat transfer capability of the winding channel configuration 5 a will be greater than that of the linear channel configuration 5 b, even though the channel wall areas are comparable. The average heat transfer capability of the winding channel configuration 6 a will also be substantially greater than that of the linear channel configuration 6 b because in addition to a higher heat transfer coefficient, the winding channel has a wall area about three times larger than that of the linear channel. These advantages would be found when used in micro-channels or larger channels, such as mini-channels. Thus, it will be readily appreciated that the winding channel configuration disclosed in the present invention aids in providing greater average heat transfer capability.

The following examples are provided as comparisons, are intended to be illustrative in nature, and are not to be considered as limiting the scope of the invention.

Example 1

To illustrate the magnitude of the heat transfer coefficient enhancement and pressure drop increase, the ratio of the heat transfer coefficient and pressure drop of a winding micro-channel with a topology similar to that depicted in FIG. 3.b to that of a straight micro-channel of equal length and cross-section were computed using ANSYS, a commercial computational fluid dynamics (CFD) software. The results of this computation are shown in FIG. 7 a. At a Reynolds number of 300, the winding micro-channel heat transfer coefficient is about three times larger than that of a linear micro-channel and the pressure drop is about 1.5 times larger. Similar results would be anticipated with a winding mini-channel design.

Example 2

A heat exchanger according to the first embodiment described above was fabricated and tested. The heat exchanger had a 60×60 mm transfer area and the heat transfer member consisted of a stack of three heat transfer layers fabricated out of 0.25 mm thick copper foil. Winding micro-channels with a width of 0.25 mm and a depth of 0.17 mm were chemically etched into one surface of the heat transfer layers. Inlet and outlet opening with a diameter of 0.75 mm were etched through the heat transfer layers. The distance between the inlet and outlet openings was 4.8 mm. The winding micro-channel topology was similar to that depicted in FIG. 3 b. It had six small-scale undulations with a pitch of 2.4 mm and amplitude of 0.75 mm and three large scale bends spanning the distance between the inlet and outlet openings.

A 20×20 mm heat exchanger with winding mini channels having a width of 1 mm and a depth of 0.65 mm was fabricated and tested. The mini channels had amplitude of 0.86 mm and a wavelength of 3 mm. The total length of the mini channels was 20 mm.

The thermal resistance of the winding micro-channel cold plate and winding mini-channel cold plate was measured as a function of the water flow rate per unit area. The measured resistance is shown in FIG. 7 b. Also shown in FIG. 7 b is the vendor provided performance for a swaged-tube cold plate commercially available from Lytron, Inc, product number CP15, a normal-flow cold plate, and the theoretically minimum resistance that can be achieved at a given flow rate per unit area. The theoretically minimum resistance corresponds to that of an ideal cold plate whose surface temperature was equal to the fluid exit temperature. As evidenced from these measurements, the winding channel cold plates (both mini and micro) can achieve a thermal resistance an order of magnitude lower than that of a prior art, exemplary swaged-tube cold plates. For the given flow rates, the thermal resistance of both the winding micro-channel and mini-channel cold plates approaches the theoretical lower bound.

It has also been found that in utilizing winding mini-channels heat transfer uniformity can be improved by varying the amplitude of the undulations of the winding channels along the length, “l” thereof, as illustrated in FIG. 7. By increasing the amplitude along the length, the heat transfer coefficient is likewise increased. As the fluid travels along the length of the winding channel, the temperature change (ΔT) is greatest at the beginning of the winding channel and decreases along the length toward the exit. By varying the amplitude to increase the heat transfer coefficient, the overall heat transfer along the length of the channels remains largely consistent from the beginning to the end of the channel. The amplitude may vary between individual undulations as illustrated in the present embodiment, or the undulations may be grouped together with two more undulations forming a group. Each group of undulations may include undulations having the same amplitude, the amplitude of the groups of undulations increasing along the length of the winding channel, moving in a direction toward the outlet openings. The heat transfer coefficient may also be increased by creating undulations having shorter wavelengths “w” along the length, “l” of the channel. For example, for any given length “l” of a channel, fewer undulations having longer wavelengths each will result in a lower heat transfer coefficient whereas for the same length channel, and same amplitude undulations, smaller wavelengths and a corresponding increase in the number of undulations will increase the heat transfer coefficient. In one embodiment, the ration of the amplitude of the undulation to the width of the channel may vary from about 0 to about 1.5, although other ratios may be utilized.

FIGS. 7 c-7 f graphically illustrate the effects of the undulations amplitude and wavelength on the mini-channel pressure drop and heat transfer coefficient. Referring initially to FIG. 7 c, a graph comparing the pressure drop of a straight channel to that of two winding mini-channels each having different amplitude undulations is shown. FIG. 7 d is a graph comparing the pressure drop of a straight channel to that of two winding mini-channels each having different wavelength undulations. FIG. 7 e is a graph comparing the heat transfer coefficient of a straight channel to that of two winding mini-channels each having different amplitude undulations while FIG. 7 f is a graph comparing the heat transfer coefficient of a straight channel to that of two winding mini-channels each having different wavelength undulations. It should be noted, that while these graphs illustrate mini-channels, they would be equally applicable to any size channels.

The substantial increase in heat transfer coefficient resulting from the winding channel geometry allows the use of mini-channels in some applications formerly requiring the use of micro-channels to achieve the desired heat transfer capacity. The mini-channel design also has the added advantage of wider channels that generally do not require filtration of the cooling liquid prior to entering the winding mini-channels. Filtration is generally required to reduce particulates that can block the small-scale micro-channel passages, impeding performance and sometimes leading to failure of the micro-channel device. As will be appreciated, the elimination of filtration that is generally required with micro-channels reduces the cost, labor and possibility of break down for the mini-channel device as compared to a similar device utilizing micro-channels.

In addition to the potential for improved heat exchanger performance discussed above, the present disclosure provides an inexpensive approach for fabricating winding channel heat exchangers to meet a wide range of applications. For example, the winding channels can be fabricated inexpensively by laser machining, chemical milling, or the like. In the chemical milling process, photosensitive resist layers are laminated to both sides of a metal foil and a photomask is employed to pattern the winding channel geometry onto the resist. After development, the resist is removed from the areas that will be etched. The winding channels may be made by etching the metal from only one side; thereby obtaining a partially etched feature that does not extend through the thickness of the layer 28. The inlet and outlet openings 32, 34 may be made by etching the metal from both sides, until all the metal is dissolved and a through feature is obtained that connects the first surface 10 a to the second surface 10 b.

The present construction also simplifies the fabrication of heat exchangers having a range of heat transfer capabilities. The heat transfer capability of the heat exchanger is proportional to the flow rate, and to maintain the same thermal effectiveness, the product of the winding channel wall area and the winding channel heat transfer coefficient must be proportional to the flow rate. In the heat exchanger of the present disclosure, this can be easily accomplished by increasing the number of layers in the heat transfer member in proportion to the required flow for the target application. In addition, the present embodiment also allows for inexpensive tailoring of the heat transfer capability over the surface of the heat exchanger. In some applications it may be desirable to provide a greater heat transfer capability (lower thermal resistance) in one area of the heat exchanger and a smaller heat transfer capability (higher thermal resistance) in another. For example, if hot spots are disposed in one area greater heat transfer capability in that area would be desired. This can be easily accomplished by using the heat exchanger of the present disclosure. For example, a flow restrictor plate 44 can be inserted between the manifold 12 and the heat transfer member 14. As illustrated in FIG. 8, the flow restrictor plate 44 includes openings 46 that allow fluid flow through the micro-channels 30 in the areas that require maximum heat transfer capacity, while the body 48 of the plate 44 selectively blocks the flow from some of the micro-channels 30 in areas that require a smaller heat transfer capacity. The pattern, size and shape of the openings 46 can be tailored according to the particular application while the winding channel design of the heat transfer member 14 remains the same. In this manner the same winding channel configuration could be used for heat exchangers having different heat transfer patterns.

Alternatively, the same effect could be achieved by grouping the winding channels so as to vary the heat transfer capability in certain areas of the heat transfer member. For example, the number of winding channels per unit area could be varied over the active area of the heat transfer member, with some groupings being denser than others so as to increase the heat transfer capability in the areas with a greater density of winding channels. Likewise, the winding channels could also be grouped according to the size of the undulations and/or bends, with similar amplitude winding channels being grouped together. In this manner, the amplitudes of the undulations could be varied between groups of winding channels over the heat transfer member so as to vary the thermal resistance over an active area of the heat transfer member.

Yet another fabrication advantage of the heat exchanger of the present disclosure is that the flow distribution and heat transfer functions are confined to different components. The heat transfer capacity depends primarily on the geometry and material properties of the heat transfer member and high thermal conductivity materials, such as copper or aluminum, need only be used in the fabrication of the heat transfer member. The manifold could be fabricated out of lower cost materials such as a temperature resistant polymer. The manifold could also be a stamping made out of a lower cost metal. Accordingly, the present invention provides for a device that can be readily tailored to a variety of needs in an inexpensive and readily achievable manner.

It will be apparent to those skilled in the art, that there are many variations in the winding channel geometry that can be used to advantage to meet the requirements of different applications. As shown in FIGS. 9 a-9 e and FIG. 10, various embodiments of designs for alternate winding channels 30 are illustrated, for example, FIGS. 9 b and 9 c illustrate non-arcuate embodiments; FIG. 9 a includes undulations 38 that change the direction of the fluid flow, but does not include bends that reverse the direction of the fluid flow; FIG. 9 d includes a single, arcuate undulation; FIG. 9 e includes winding micro-channels 30 are etched on both sides of layer 28 to increase the micro-channel density; and FIG. 10 where each winding micro-channel 30 is separated from every other micro-channel 30. All of these designs are variations of the first embodiment and operate under the same principals.

It will be understood that various modifications may be made to the embodiments disclosed herein. For example, the dimensions and geometric shapes may be modified, as would be known to those of skill in the art. In addition, the winding channels may find use in normal flow cold plates as well as parallel flow cold plates in which case the directional examples would be modified. In addition, the number and size of the small amplitude undulations and reverse bends can be varied depending upon the application, and some applications may only have bends that reverse the direction of the fluid flow, while others may only have undulations that change the direction of the fluid flow and some may have both. In addition, the winding channel's axis may remain non-linear along only a portion of the length of the winding channel. Likewise, the examples provided are not to be construed as limiting, but as projected outcomes of exemplary embodiments. Therefore, the above description should not be construed as limiting, but merely as exemplifications of preferred embodiments. Those skilled in the art will envision other modifications within the scope, spirit and intent of the invention. 

1. A heat exchanger comprising: a heat transfer member including at least one heat transfer layer; one or more inlet openings disposed in the heat transfer member; one or more outlet openings disposed in the heat transfer member; at least one winding channel disposed in each of the at least one heat transfer layers and constructed and arranged to carry a fluid, the at least one winding channel having a length and a non-linear flow axis, the non-linear flow axis defining a non-linear path between the one or more inlet openings and the one or more outlet openings; at least one undulation having an amplitude constructed and arranged to increase the heat transfer coefficient of the fluid as it passes there through and further being constructed and arranged to change the direction of the flow of the fluid as it travels along the non-linear flow path; wherein during use fluid flows through the one or more inlet opening in the heat transfer member, into the at least one winding channel, and flows into the at least one undulation, the amplitude of the undulation increasing the heat transfer coefficient as the fluid passes there through, and wherein the fluid continues to move toward the outlet opening after changing direction and passing through the at least one undulation.
 2. The heat exchanger of claim 1, wherein the at least one undulation comprises a plurality of undulations, and wherein the amplitude of the undulations are constant along the length of the at least one winding channel.
 3. The heat exchanger of claim 1, wherein the at least one undulation comprises a plurality of undulations, and wherein the amplitude of the undulations vary along the length of the at least one winding channel.
 4. The heat exchanger of claim 3, wherein the amplitude of the plurality undulations increases along the length of the at least one winding channel moving in a direction from the one or more inlet openings toward the one or more outlet openings.
 5. The heat exchanger of claim 1, wherein the at least one winding channel comprises a plurality of winding channels, the plurality of winding channels each having one or more undulations, and wherein winding channels having undulations of similar amplitude are grouped together, the amplitudes of the undulations varying between groups of winding channels over the heat transfer member so as to vary the thermal resistance over an active area of the heat transfer member.
 6. The heat exchanger of claim 1, further comprising a manifold including an inlet port and an outlet port constructed and arranged to distribute fluid to and collect fluid from the heat transfer member.
 7. The heat exchanger of claim 1, wherein the at least one winding channel comprises mini-channels.
 8. The heat exchanger of claim 1, wherein the at least one winding channel comprises micro-channels.
 9. The heat exchanger of claim 1, wherein the non-linear flow path of the at least one winding channel includes an inlet side adjacent a corresponding inlet opening and an outlet side adjacent a corresponding outlet opening, the non-linear path further including at least one pair of bends, each pair having: a) a first bend constructed and arranged to reverse the direction of the flow of the fluid as it travels between the corresponding inlet opening and the corresponding outlet opening such that the fluid flows toward the inlet side after passing through the first bend; b) a second bend constructed and arranged to reverse the direction of the flow of the fluid as it travels between the corresponding first opening and the corresponding second opening such that the fluid flows toward the outlet side after passing through the second bend; and wherein during use fluid flows from the manifold, through the corresponding inlet opening in the heat transfer member, into the at least one winding channel and travels along the non-linear path toward the outlet side of the non-linear channel and flows into the first bend which reverses the direction of the fluid flow toward the inlet side of the channel, the fluid thereafter flowing into the second bend which reverses the direction of the fluid flow toward the outlet side of the channel, the flow of fluid traveling along the non-linear path to the outlet opening.
 10. The heat exchanger of claim 9, wherein the first bend, the second bend and the at least one undulation each have an arcuate shape.
 11. The heat exchanger of claim 1, wherein the depth of the at least one winding channel is substantially equal to the width of the at least one winding channel.
 12. The heat exchanger of claim 1, wherein the at least one heat transfer layer comprises a bonded stack of at least two laminations.
 13. The heat exchanger of claim 1, wherein the at least one heat transfer layer comprises a single heat transfer layer.
 14. A heat exchanger comprising: a heat transfer member including at least one heat transfer layer, the at least one heat transfer layer having a first surface and a second surface and including a thickness extending between the first surface and the second surface; a manifold including an inlet port and an outlet port constructed and arranged to distribute fluid to and collect fluid from the heat transfer member; one or more inlet openings disposed in the heat transfer member and in fluid communication with the manifold; one or more outlet openings disposed in the heat transfer member and in fluid communication with the manifold; at least one winding channel disposed in each of the at least one heat transfer layers, the at least one winding channel having a non-linear flow axis, the non-linear flow axis defining a non-linear path between the one or more inlet openings and the one or more outlet openings, the non-linear flow path having an inlet side adjacent a corresponding inlet opening and an outlet side adjacent a corresponding outlet opening, the non-linear path further including at least one pair of bends, each pair having: a) a first bend constructed and arranged to reverse the direction of the flow of the fluid as it travels between the corresponding inlet opening and the corresponding outlet opening such that the fluid flows toward the inlet side after passing through the first bend; b) a second bend constructed and arranged to reverse the direction of the flow of the fluid as it travels between the corresponding first opening and the corresponding second opening such that the fluid flows toward the outlet side after passing through the second bend; wherein during use fluid flows from the manifold, through the corresponding inlet opening in the heat transfer member, into the at least one winding channel and travels along the non-linear path toward the outlet side of the non-linear channel and flows into a first bend which reverses the direction of the fluid flow toward the inlet side of the channel, the fluid thereafter flowing into the second bend which reverses the direction of the fluid flow toward the outlet side of the channel, the flow of fluid traveling along the non-linear path to the outlet opening.
 15. The heat exchanger of claim 14, wherein the at least one winding channel comprises mini-channels.
 16. The heat exchanger of claim 14, wherein the at least one winding channel comprises micro-channels.
 17. The heat exchanger of claim 14, wherein reversing the direction of the fluid flow toward the inlet side and back toward the outlet side of the non-linear path results in substantially uniform thermal resistance throughout the heat transfer member.
 18. The heat exchanger of claim 14, further comprising at least one undulation constructed and arranged to change the direction of the flow of the fluid as it travels along the non-linear flow path, without reversing the direction of the flow of the fluid, wherein if the fluid is moving toward the outlet side before passing through the at least one undulation, the fluid continues to move toward the outlet side after passing through the at least one undulation and wherein if the fluid is moving toward the inlet side before passing through the at least one undulation, the fluid continues to move toward the inlet side after passing through the at least one undulation.
 19. The heat exchanger of claim 18, wherein the first bend, the second bend and the at least one undulation each have an arcuate shape.
 20. The heat exchanger of claim 14, wherein the depth of the at least one winding channels is substantially equal to the width of the winding channel.
 21. The heat exchanger of claim 14, wherein the at least one heat transfer layer comprises a bonded stack of at least two laminations.
 22. The heat exchanger of claim 14, wherein the at least one heat transfer layer comprises a single heat transfer layer.
 23. The heat exchanger of claim 14, further comprising a flow restrictor plate disposed between the manifold and the heat transfer member, the flow restrictor plate including a body having a plurality of openings disposed there through that are configured, dimensioned and positioned to vary the flow to the winding channels according to the heat transfer requirements.
 24. The heat exchanger of claim 23, wherein the body of the flow restrictor plate is constructed and arranged to selectively block the fluid flow to the winding channels in sections of the heat transfer member
 25. The heat exchanger of claim 14, wherein the at least one winding channel comprises a plurality of winding channels, the plurality of winding channels being disposed over an active area of the heat transfer member in grouping of two or more, the number of winding channels per unit area varying between groups so as to vary the thermal resistance over the active area of the heat transfer member. 